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CFD and Frictional Torque Analysis of
Hydrodynamic Journal Bearing
Abhinav Maurya Department of Mechanical Engineering
Delhi Technological University
New Delhi, India
Abhay Kumar Department of Mechanical Engineering
Delhi Technological University
New Delhi, India
Aman Dev Punia Department of Mechanical Engineering
Delhi Technological University
New Delhi
Abstract— This paper aims at CFD and frictional torque
analysis of hydrodynamic journal bearing. Numerous
experiments were performed on bronze bearing and two journals
of different materials, mild steel and bronze for the frictional
torque analysis using SAE10W40 lubricating oil on specially
designed test rig. Graphs for frictional torque and coefficient of
friction were plotted from results of experimental data and
theoretical calculation. Results of both journals were compared
for frictional torque and coefficient of friction was calculated
experimentally and theoretically (using the Petroff’s equation). It
was observed that frictional torque increases as the shaft speed
increase and becomes almost constant as the shaft approaches
extreme speed.
Keywords — Hydrodynamic Journal Bearing, Frictional
Torque Analysis, Stribeck Curve, Reynold’s Equation
I. INTRODUCTION
Bearing, especially hydrodynamic bearings are used to support
rotating shaft in numerous industries. Journal bearing
generally, operates with hydrodynamic lubrication. It has a
shaft or journal rotating in a bearing. In the bearing, the shaft
rotates with a layer of lubricant separating the two parts [Note:
journal is the part of shaft lying inside the bearing]. It is used
when the load is light and the motion is continuous. Since the
journal bearing has no rolling element as in other bearings, it
is also known as radial load sliding contact bearing.
Hydrodynamic journal bearing (fig. 1) is a very simple yet
critical component as there are numerous parameters involved
in its proper working. Journal bearing works by creating a
pressure difference by moving fluid from converging to
diverging by wedging action of fluid due to moving body parts
because of which hydrodynamic journal bearing surface
velocity and gap generation varies. Friction in fluid film causes
frictional force and wear which reduces performance
efficiency and increases maintenance cost. In hydrodynamic
journal bearing, there is metal-to-metal contact during start-up.
However, during running condition, there is fluid film between
the two moving surfaces resulting less power loss and
increased efficiency.
Failure in the Journal Bearings occurs due to improper
lubrication as it will cause more wear rate and thermal
expansion. Lubricant keeps the moving parts’ surface away
Fig. 1: Journal Bearing
from each other, acts as cooling agent and a medium to transfer
additives to the machining area. Lubrication in the journal
bearings is usually done by grooving that distributes and
supplies the lubricant (placement of grooves is away from the
load zone). Load, surface roughness, viscosity, speed, and
clearances determine the correct oil film thickness for the
specific application. It is often used in industrial machines that
require high horsepower and high loads like turbines and
pumps.
A. Basic Concept of Hydrodynamic Journal Bearing
The basic concept in hydrodynamic journal bearing is the
lubricant action and its flow in between journal and the bearing,
i.e. how the shaft is lifted by the lubricant to avoid metal-to-
metal contact (see fig. 2). The lifting of the shaft is due to the
wedging action (also called dynamic action or convergent
action).
Steps in wedge development (Bearing speed is denoted by N):
i. N1 = 0 rpm, i.e. idle condition, the shaft rests on the bearing
surface with normal reaction equal to the weight of the shaft.
ii. N2 > N1, i.e. the working condition, even though the shaft has
started rotating, the speed is very low. Thus, there is a constant
contact between the bearing and the journal surfaces. At this
speed, the pressure is maximum at the point of contact.
iii. N3 > N1, i.e. at high speed (increasing rate), the shaft starts
lifting off. As the shaft speed increases, due to the friction
between the shaft and the lubricant, the lubricant enters under
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Fig. 2: Basic Development of an Oil Wedge
the shaft through converging action. Thus, lifting the shaft and
the maximum pressure is at the location of minimum film
thickness.
iv. N4 >>> N1 i.e. at high speed (constant or stable speed), the
shaft is stable and the lubricant film thickness between the
bearing and the journal in this situation is termed as minimum
film thickness.
B. Frictional Torque
When the shaft is in stationary condition (absence of friction),
the reaction offered by bearing is inline with a line of action of
the load acting from the journal. When the shaft is in motion
(presence of friction), the resultant R is deviated by distance x
from the line of action of the load acting from the journal. A
circle is drawn from the center of the shaft by taking radius x
is known as friction circle.
Frictional Torque is given by,
Tf = µWr
where,
µ = Coefficient of friction
W = Load acting on bearing (N)
r = radius of bearing
The coefficient of friction, µ, can also be calculated using
Petroff’s Equation.
Petroff’s Equation: µ = 2𝜋2 (𝑍𝑛
𝑃) (
𝐷
𝐶) + 𝐾
where, 𝑍𝑛
𝑃 = Bearing Characteristics Number
Z = Lubricant Viscosity (Pa-s)
n = Rotation per second
P = Pressure (Pa)
D = Bearing Diameter (mm)
C = Diametrical Clearance (mm)
K = Leakage Factor
K = 0.002 when 𝐿
𝐷< 0.75
K = 0.003 when 𝐿
𝐷≥ 0.75
C. Stribeck Curve
Stribeck curve (see fig.3) is a curve plotted between the
coefficient of friction against the bearing parameter (ZN/P),
Fig. 3: Stribeck Curve
lubricant viscosity, N is rotating speed per second and P is
bearing pressure. ZN/P is also known as bearing characteristics
number (BCN).
• Unstable Region: Boundary lubrication and mixed
lubrication regime lie in this region. In the boundary
lubrication, there is a constant metal-to-metal contact hence
the coefficient of friction is high. However, as we proceed
towards the mixed lubrication, the coefficient of friction
starts decreasing as the film thickness increases. In addition,
due to the heat generated, the lubricant viscosity starts
decreasing which reduces the effect of wedging action.
Combining all the factors it can be noticed that as the BCN
decreases, viscosity decreases and hence the coefficient of
friction increases.
• Stable Region: Hydrodynamic lubrication regime lies in this
region. As the journal speed increases, BCN increases and in
addition to that heat generation also increases. In this case,
amount of heat generation dominates, thus causing an
increase in temperature of lubricant. Therefore, decreasing
the lubricant viscosity and hence increasing the coefficient
of friction The point where the coefficient of friction is
minimum is called bearing modulus.
In the unstable region, the power loss is dependent on the load
applied. On the other hand, in the stable region, the power loss
is dependent on viscosity instead of load applied.
D. Reynold’s Equation
Reynold’s equation is generally used in modelling of the
hydrodynamic lubrication regime and is used to aid
calculations. Pressure distribution in the lubricant can be
expressed as a function of journal speed, bearing geometry, oil
clearance and lubricant viscosity using Reynolds equation.
There are certain assumptions made in hydrodynamic models:
• Negligible body forces
• Constant pressure throughout the film thickness
• No slip at the boundaries
• The lubricant is a Newtonian fluid (not true for polymeric
oils)
• Flow is laminar (not true in large bearings)
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• Fluid inertia is negligible
• Fluid density is constant (valid for liquid lubrication, not
gas lubrication)
• Viscosity is constant throughout the film (not exactly true
but it simplifies the problem)
Reynold’s equation: 𝜕
𝜕𝑥(
ℎ3
µ
𝜕𝑝
𝜕𝑥) +
𝜕
𝜕𝑧(
ℎ3
µ
𝜕𝑝
𝜕𝑥) = 6U
𝜕ℎ
𝜕𝑥
where:
h – local oil film thickness,
η – dynamic viscosity of the oil,
p – local oil film pressure,
U – linear velocity of the journal,
x - circumferential direction.
z - longitudinal direction.
Close form solution of Reynolds equation can’t be obtained
therefore finite element method is used to solve it. Two
analytical solutions of Reynolds equation exist only for above-
mentioned assumptions. The two solutions are Sommerfeld
and Ocvirk solutions, are applicable only in the region of
positive pressure
• Sommerfeld Solution:
𝑃 = Ƞ𝑈𝑟
𝑐𝑟2 (
6ɛ(2+ ɛ 𝑐𝑜𝑠𝛳)𝑠𝑖𝑛𝛳
(2+ɛ2)(1+ɛ𝑐𝑜𝑠𝛳)2) + 𝑃𝑜
The equation is solved with the assumption that there is no
lubricant flow in the axial direction (infinitely long bearing
assumption).
• Ocvirk Solution: Ocvirk solution for infinitely short bearing
assumption neglects circumferential pressure gradients (first
term of Reynolds equation)
𝑃 = Ƞ𝑈𝑟
𝑐𝑟2 (
𝐵2
4− 𝑧2) (
3ɛ𝑠𝑖𝑛𝛳
(1+ɛ𝑐𝑜𝑠𝛳)3) + 𝑃𝑜
𝑃 = 𝑃𝑐𝑣
𝐷2𝑆𝑜(𝐵2 − 4𝑧2) (
3ɛ𝑠𝑖𝑛𝛳
(1+ɛ𝑐𝑜𝑠𝛳)3) + 𝑃𝑜
where,
Cr: radial clearance
ε: eccentricity ratio
B: bearing length,
Po: cavitation pressure,
So: Sommerfeld number.
The Sommerfeld number remains constant for a given bearing
and is used to correlate working condition of the different
machine working with the same bearing. Sommerfeld number
is analogous to the human fingerprint, i.e, every bearing has its
own unique Sommerfeld number. It is given by
Sommerfeld Number: So = 𝑍𝑁
𝑃(
𝐷
𝐶)2
II. CFD ANALYSIS
In this research work, the pressure distribution in the
lubricating fluid regime was solved in which the inlet of fluid
was provided through the mid-section and the outlet of the
fluid was taken at the end walls of the bearing at both sides.
Further, the result obtained was used to solve for the
deformation, stresses, strain etc. The CFD analysis was applied
to mathematically check whether the pressure distribution in
the fluid film was same as experimental. Different simplifying
assumptions, based on experimental observations were used to
solve the characteristics of bearing. The assumption was made
that the viscosity of fluid film remains same and does not vary
with the temperature. The energy equation was applied with
the assumption that the regime of flow to be laminar. The
bearing load was applied along with the effect of fluid pressure
contour on the bearing was applied to calculate the stresses,
strain, strain energy, deformation of the bearing. The solution
was done with the use of CFD and ANSYS Fluent software.
A. Geometrical modelling and Meshing
The bearing material taken is Bronze (86% copper, 9.5% tin,
2.5% lead, and 2% zinc). However, the materials used for the
journals were Bronze (86% copper, 9.5% tin, 2.5% lead, and
2% zinc) and Mild steel (0.05%-0.25% carbon and up to 0.4%
manganese, rest Iron). The design of journal bearing was made
with the help of ANSYS 15.0 (see fig. 4), the dimensions of
which are shown in table 1.
For the fluent analysis, the journal and bearing were
suppressed and the fluid film between them having a thickness
of 27 microns was considered. The mesh was generated over
the fluid film (fig. 5). The named selection was created where
the inlet was provided in the mid-section of the fluid film and
the outlet was marked at the end of the plane surface.
Fig 4: 3D Model of the journal bearing in ANSYS
TABLE 1: DIMENSIONS OF JOURNAL BEARING MODEL USED IN
ANSYS
S.No Part Detail Range
1. Shaft Diameter 22+0.005mm
2. Journal Outer Diameter 40-0.005/0.006mm
3. Journal Inner Diameter 22.035mm
4. Bearing Inner Diameter 40+0.050/0.06mm
5. Bearing width 40.04mm
6. Radial Clearance 0.027mm
7. L/D ratio 1
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Fig 5: Meshing of fluid film showing named selection in ANSYS 15.0
B. Solution
For solution, the laminar flow was selected and the energy
equation was turned on. The material of the fluid was then
input as given below. The properties of the lubricant used
(SAE10W40) was entered as the data mentioned in Table 2.
After this, the boundary conditions were inserted in which the
velocity inlet criteria was followed. The inlet velocity was
taken as 10m/s and the inlet pressure was taken as 101325
pascals. The temperature of the fluid inlet was taken to be
around 320K. After running iterations and solving the pressure
distribution on the journal bearing was found as pressure
contour.
TABLE 2: LUBRICANT PROPERTIES
S.No. Parameter Value
1. Journal Radius 40 mm
2. Radial Clearance, C 27µm
3. Lubricant Viscosity, µ 0.079 kg/ms
4. Lubricant Density 845 kg/m3
5. Lubricant specific heat, Cp 1901.07J/kgoC
6. Lubricant thermal
conductivity
0.1242
C. Analysis of Journal Bearing
In the static analysis, the mesh was generated around the
bearing (fig. 6), then the load was imported from the previous
contour at the inner surface of the cylinder. The bearing load
was also inserted in the component form (fig. 7). The solution
was done and various results for the equivalent stress, strain,
deformation and strain energy was calculated. The bearing
material properties considered are shown in table 3.
Fig 6: Mesh generation on bearing
TABLE 3: BEARING MATERIAL PROPERTIES
S.No. Property Value
1 Density 8719 kg/m3
2 Young’s Modulus 103 GPa
3 Poisson’ Ratio 0.34
4 Bulk Modulus 116 GPa
5 Shear Modulus 40 GPa
Fig 7: Pressure imported on bearing
III. MODELLING BASED ON EXPERIMENTAL TEST
RIG
A. Main Designing Frame
The Journal Bearing Rig TR-660 (fig. 8-10) was used to
demonstrate the frictional torque and pressure distribution
inside bearing. Table 4 gives the mechanical specifications of
the test rig. The journal was mounted horizontally on a shaft
supported on a self-aligning bearing and a motor with timer
bolt rotates the shaft. A leaded bronze, flawless bearing freely
slides over the journal and as it rotates radial load was applied
on bearing by pulling it upwards against journal by a loading
lever. Ten sensors were fixed on the circumference of bearing
with their terminal ends ending in a junction box. The journal
was driven by a belt and two-step pulley arrangement and
speed required was set on software. Radial load and journal
speed were varied to suit the test conditions. The testing oil
used for the experiment was SAE 10W40.
A ground shaft was supported on two self-aligned bearings on
either end of the oil sump. The oil sump was wide enough to
house the bearing with all sensors with the connector inside it.
To attain a low speed of 40rpm and max speed of 8000rpm,
two-step pulleys having a ratio of 1:4 and 4:1 ratio were fixed
on shaft and motor ends. A motor having a base speed of
1435rpm was mounted over base plate drives the shaft.
The loading lever having load ratio 1:5 was fixed on a frame,
the frame height was kept sufficiently high to minimize the
reduction of radial load due to the inclination of loading lever
when max weight was placed. The loading lever was pivoted
to a vertical frame, smaller and of the lever was connected to
bearing top through a wire rope, on either end of wire spherical
eye bolt was fixed to give free angular movement.
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Fig. 8: TR660 Journal Bearing Test Rig
Fig 9: Loading System
Fig. 10: Front view and side view of the journal bearing test rig
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TABLE 4: MECHANICAL SPECIFICATIONS OF THE TEST RIG
S.No Part Detail Range
1. Shaft Diameter 22+0.005mm
2. Journal Outer Diameter 40-
0.005/0.006mm
3. Journal Inner Diameter 22.035mm
4. Bearing Inner Diameter 40+0.050/0.06mm
5. Bearing Outer Diameter 95mm
6. Bearing width 40.04mm
7. Radial Clearance 0.027mm
8. L/D ratio 1
9. C/R Ratio 0.00135
10. Max. Load 3000N
11. Loading Ratio 1:5
12. Frictional Torque Distance 51mm
13. Spindle Speed (min.) 40 rpm
14. Spindle speed (max.) 8000 rpm
15. Lubrication Unit (Oil Pressure) 5-10 bar
16. Lubrication Unit (Tank Capacity) 25 L
17. Lubrication Unit (Size l*b*h) 370 x 290 x
300mm
18. Lubrication Unit (make) Cenlube System
19. Oil Hole (external surface) 13mm
20. Oil Hole (internal surface) 6mm
B. Measuring system – Frictional Torque Sensor
The measuring system measures frictional torque and records
the data. Table 5 shows the specification of the frictional
torque sensor. This system consists of computer, signal
processing modules and measuring devices as its main
components. A beam type load cell 30kg capacity was fixed
on to the oil sump; the frictional arm touches the oil inlet
researching from the bearing. During the journal rotation, the
force exerted was transferred to the load cell that was measured
and displayed on the computer screen. The torque distance was
51mm.
TABLE 5: SENSOR SPECIFICATIONS
Part Detail Specification
1. Speed
• Sensor Output From Drive
• Range 40-8000rpm
• Least
Count
1rpm
• Accuracy (1±2% measured speed)rpm
2. Frictional
Torque
• Sensor Beam Type load cell make: Sensortronics, cap
30 kg
• Range 15Nm
• Least
Count
0.01Nm
• Accuracy (0.1±1% measured value) Nm
IV. EXPERIMENTAL ANALYSIS
A. Procedure
The experiment was conducted for two different materials for
the journal, mild steel and bronze with the bearing made of
bronze. The following procedure was for the journal made of
mild steel.
1. Selection of test bearing: The test bearing made of Bronze,
with an inner diameter of 40.06mm was used. The bearing was
a square bearing i.e. the l/d ratio of the bearing was one.
2. Selection of journal: Journal was selected according to the
c/r ratio required. The journal selected has c/r =0.00135 and
the outer diameter of the journal was 40.005mm.
3. Machine preparation for testing: Power input cable was
connected to 220V, 50Hz, 3 phase supply. Then the machine
was turned on by switching on the MCB on machine side
panel, followed by switching on the lubrication motor.
4. Preparation of bearing for test:
a. Split housing on either end of the self-aligned bearing was
loosened and removed.
b. A bush holder and the journal was inserted into the shaft
and slide the other bush holder.
c. The bearing was placed inside the oil sump at middle and
the bracket was inserted over the bearing.
d. Shaft with the journal was pushed through the test bearing
and was provided with support on the bearing.
e. Lock nuts on either side of the self-aligned bearing were
tightened.
f. Driven pulley was tightened onto the shaft, followed by
loosening the motor position and aligning the driver
pulley in line with driven pulley.
g. The belt was passed over the pulleys and the nuts were
tightened using four screws to lock the motor position.
h. The weight of 1kg-wt was placed over the loading pan,
followed by pushing the bracket upwards and moving the
bearing inside the bracket.
i. The bearing inside the bracket was positioned properly and
the journal was pushed till it was inside the bearing.
j. Finally, the journal was locked in place by clamping the
bush holder.
5. Setting the computer for test:
a. Winducom 2008 software was opened and was ran in
continuous mode for operating software.
b. The acquiring screen was opened by clicking ACQUIRE
button.
c. The file name was entered in the remarks column.
d. Required journal speed was set by entering the value and
the applied load was entered in load column.
e. The START button was clicked followed by the
INITIALISE button to bring pressure button to zero and
ZERO button to bring the frictional torque to zero.
6. Testing:
a. Lube unit was turned on and the oil was allowed to run
for some time.
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b. It was ensured that the oil was flowing from either end of
the bearing.
c. 500rpm was selected from the speed range: 500rpm or
8000rpm, and the required speed (initially 50rpm) was
entered in the respective window (the actual bearing
speed was 37 times the entered speed)
d. Clicked on RUN MOTOR to start shaft rotation.
e. After the set speed was reached, dead weight was placed
on loading pan.
f. The frictional torque reading was noted for the given load
after 10 seconds.
g. Clicked on MOTOR STOP.
h. The same process was repeated for various loads (10N,
20N, 30N, 40N 50N) and for various speeds (entered
value were 50rpm, 60rpm, 70rpm, 80rpm, 90rpm).
Similarly, the experiment was conducted for the journal made
of bronze by replacing the journal made of mild steel and the
readings were noted.
B. Precautions taken during the testing
a. To prevent the machine damage, the bearing temperature
was not allowed to exceed 125oC.
b. The lubricant was allowed to be cooled down to prevent
viscosity changes.
c. After five consecutive readings, the machine was turned
off to allow the bearing and the lubricant to be cooled
down.
V. RESULTS AND DISCUSSION
A. Results of CFD analysis
The pressure contour generated on the fluid film depicts that
the pressure distribution varies greatly around the journal. The
pressure was varying along the circumference of the oil film
and the pressure being maximum on the lower side of the oil
film, in this region the film was convergent. This can be related
to the wedging action of the fluid film in which the pressure
was maximum near the minimum film thickness region (see
fig. 11). Similar results were also obtained for the stress using
static analysis (fig. 12). This was validated with experimental
as well as analytical results.
The various other results related to the deformation and stress
in the bearing due to the film pressure and bearing load was
also calculated using the computational fluid dynamics. The
deformation in the bearing can be related to the pressure wave
generated around the circumference of the bearing due to
conversing and diverging phenomena respectively in the
direction of rotation. This result showed that net deformation
was maximum at lower left part of the bearing when the
rotation was clockwise during the startup (see fig. 13).
B. Results of Frictional torque
The frictional torque was calculated for the various load and
rotation of the shaft. The shaft and journal were made of
stainless steel, the bearing was made of bronze. On this setup
Fig 11: Pressure contour of the oil film using ANSYS Fluent
Fig 12: Stress in the bearing using ANSYS static analysis
Fig 13: Total deformation in bearing using ANSYS structural analysis
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the load was applied as 100N, 150N and 250N and the rotation
speed were varied as 1850rpm, 2200rpm, 2600rpm, 3000rpm,
3350rpm. The load ratio provided was of 1:5 and the frictional
torque distance was 51 mm.
The frictional torque was calculated for the various load and
the graph was plotted was plotted against shaft speed (in rpm),
where the frictional torque (in Nm) was plotted on the y-axis
and shaft speed (in rpm) was plotted on the x-axis for various
loads as earlier mentioned (see fig. 14). ). It was observed that
the frictional torque increases with increase in shaft speed but
become almost constant at high speeds.
C. Results of the Coefficient of friction
Variation of co-efficient of friction was plotted against shaft
speed (in rpm), where the coefficient of friction was plotted on
the y-axis and shaft speed was plotted on the x-axis for various
loads. For the bearing load of 250N, the coefficient of friction
was calculated using experimental data obtained through
specifically designed test rig machine. In addition, for the same
load of 250N, the theoretical value of the coefficient of friction
was calculated using the Petroff’s equation for various speeds
of the shaft provided.
The results obtained were then plotted to show the variation of
coefficient of frictional with varying speed of shaft for
experimental and theoretical results (see fig. 15).
The experimental and theoretical result shows that the
coefficient of friction increases as the speed of rotation of shaft
increases. It was observed from the experimental data that the
coefficient of friction between the journal and the bearing was
varied from 0.2 to 0.35 and the average calculated value was
0.301. However, from the theoretical data, it was observed that
the value of the coefficient of friction was varying from 0.25-
0.31 and the average calculated value was 0.283.
D. Effect of the Material of Journal on Frictional Torque
The effect of journal material on the frictional torque was
calculated employing the two different materials, one was mild
steel and another was bronze alloy. The result of frictional
torque was obtained for both the material and graph was
plotted to represent them graphically (see fig. 16). The load
applied was 250N and the rotating speed of the shaft was
varied from 1850 to 3350 rpm.
For two type of shaft material (steel and bronze) variation of
frictional torque was plotted against shaft rpm in which
frictional torque was on y-axis and shaft speed in rpm on the
x-axis at 250N load. The graph plotted to depict that the
frictional torque in a journal made of bronze was less than that
of bronze for the same speed of the shaft.
Fig 14: Variation of Frictional Torque with Speed for the various loads
Fig 15: Graph representing the variation of coefficient of friction
theoretically and experimentally with shaft speed at the applied load of 250N
Fig 16: Variation of Frictional Torque with Speed for a load of 250N for the two journal materials (mild steel and bronze)
1850 2200 2600 3000 3350
FRIC
TIO
NA
L TO
RQ
UE
RPM
Frictional Torque
100 N 150 N 250 N
1850 2200 2600 3000 3350
co e
ffic
ien
t o
f fr
icti
on
RPM
Coefficient of Friction
theoretical(petroff) experimental
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VI. CONCLUSION
a) The computational fluid dynamics of pressure distribution
around the fluid film showed that the pressure variation along
the circumference is varying greatly. The pressure being
maximum near the minimum film thickness region. The result
can be further used to design bearing in such a way that it could
bear the stress generated in various region. The pressure
contour generated could be related with the mathematical
results obtained by solving the Reynold’s equation. This CFD
analysis can be further proved with the help of experimental
results obtained by specially designed test rigs.
b) The frictional torque value obtained by the experimental
results on the bearing showed that the torque increases with
increasing load at the given spindle speed. the lubricating oil
SAE 10-W40 was used and It showed that with increase in
spindle speed at given load the frictional torque increases for
low range speed. The graph generated also showed that the
frictional torque becomes almost constant above 3000 RPM.
c) The coefficient of friction calculated experimentally and
theoretically using Petroff’s equations at 250 N load. The
experimental value of coefficient of friction is about 0.301 and
the theoretical value calculated using the various parameters is
0.273. The percentage error in the result is around 10.2%. This
showed that experimental analysis can be related with the
mathematical equations for hydrodynamic journal bearing.
d) The frictional torque was calculated for different materials
of same dimensions. A steel journal and another bronze alloy
journal was installed on the test rig to calculate the frictional
torque having SAE 10-W40 as lubricating oil. The result
showed that frictional torque in bronze is somewhat less than
frictional torque in journal made of mild steel. This can be
related with the fact that bronze has 2.5% lead in its
composition. As lead is a self-lubricating metal it reduces the
friction between the bearing and journal. Thus, bronze bearing
can be used instead of mild steel as it experiences less friction,
but it also experiences more wear than mild steel journal as it
a soft metal. Lead is also injurious for health as it can result in
pollution of ground and can cause hazards to human.
VII. FUTURE SCOPE
1. Pressure measurement can be done over the bearing surface
using experimental analysis.
2. Different lubrication regimes can be studied to evaluate
frictional torque.
3. Different lubricants can be studied to find the frictional
torque between the bearing and the journal
4. The effect of temperature on viscosity can be studied to
examine the dependence of frictional torque on viscosity
5. Different materials can be used for journal to further reduce
the frictional torque.
6. Use of organic materials for the journal and the bearing for
the study of frictional torque
7. Use of nanomaterials as solid lubricant to reduce the
frictional torque
8. Experimental fluid film thickness can be measured at
different locations on the periphery of the journal.
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